Modular crankshaft and connecting rod bearing assembly

ABSTRACT

An engine which employs a cam follower mechanism to reduce wear and reduce the size of an assembled engine. The cam follower mechanism utilizes guide rails located to reduce side thrust on the valve stem. The engine employs a high speed quill shaft to synchronize independent cam shafts existing in each of a plurality of interconnected engines. The engine is assembled using a single size fastener to provide a uniform stress gradient within the engine. The engines are interconnected utilizing O-ring seals. The engine provides a piston crown utilizing a connecting rod directly connected to the bottom surface of the piston crown. The piston crown is stabilized along the longitudinal cylinder axis by a rail guide. Connecting rods are provided which require less than one hundred eighty degrees (180°) circumference of a crankshaft pin for support so that a plurality of connecting rods can be associated with a single crankshaft pin. A tabbed bearing fits under the plurality of connecting rods to provide lubrication between the connecting rods and the crankshaft pin. Connecting rods are held to the crankshaft pin by a circular retaining ring. The engine provides a separate cylinder head and cylinder which are attached via a circular deformable retaining band to form a metal to metal seal. The engine provides an independent lubrication system in each engine. Coolant or lubricant is provided to each engine in parallel so that the temperature of the coolant entering each engine is the same. A large diameter modular crankshaft is provided.

This is a divisional of application Ser. No. 08/387,152, filed Feb. 10,1995; which was a divisional of application Ser. No. 08/167,193, filedDec. 13, 1993, now U.S. Pat. No. 5,429,080.

BACKGROUND

1. Field of the Invention

The present invention relates to internal combustion reciprocatingengines and in particular to a reduced size internal combustionreciprocating engine of which a plurality can be interconnected to forma larger engine.

2. Description of the Related Art

Internal combustion reciprocating engines have been known for over acentury. The internal combustion reciprocating engine has beenmanufactured in numerous configurations over the years. These enginesare utilized in automobiles, air planes and water craft. An importantconsideration in each of these applications is the size and weight ofthe engine. There is a trade off between the structural integrity ordurability of an engine and the size and weight of the engine. Enginemanufacturers design overly massive engine pans to increase thedurability and useful life of an engine. Utilization of massive engineparts, however, increases the weight and size of the engine and canactually increase engine wear by increasing the dynamic weight of themoving parts in the engine. Thus there is a need for a reduced weightand size engine that is durable.

Some engine manufacturers have apparently built engines byinterconnecting a set of smaller engines or modular engines. Modularengines are known in the prior art as evidenced by the Voorhies patent,U.S. Pat. No. 2,491,630, entitled "An Engine Constructed of SectionsBolted Together Along the Vertical Plane to Forman Entire Head Block andCrankcase Thereof," issued on Dec. 20, 1949. Voorhies patented aninternal combustion engine constructed from a series of engine modules.The Voorhies engine however suffers the same inadequacies as otherconventional engine designs.

Some of the problems presented by typical engine designs are discussedbelow.

Cam Followers

Typical cam follower mechanisms act as an intermediary between a camshaft lobe and a valve stem. Cam followers compensate for rotating camlobes side thrust. Lobes assert a composite thrust containing both ahorizontal (side thrust) and vertical (downward thrust) component. Thecam followers absorbs some of the side thrust. Any portion of thishorizontal thrust component which is asserted on the valve stemincreases wear on the valve stem and valve stem guide in which the valvestem slides. The horizontal and vertical components are asserted uponthe cam follower by the rotating cam lobe. The cam lobe rotates,depresses the cam follower mechanism, which in turn depresses the valvestem. Typically a portion of the side thrust component is notcompensated for by the cam follower. This side thrust is asserted on thevalve stem which increases wear on the valve stem and the valve stemguide.

Typical engine designs typically provide minimal lubrication between thevalve stem and the valve stem guide. Inadequate lubrication exacerbatesthe effect of wear caused by the side thrust asserted on a valve stem bythe typical cam follower mechanism. Typically, engine designers utilizelong valve stems to provide a relatively long longitudinal dimension, orhigh aspect ratio of length to width, in order to achieve stability of avalve stem along its axial length.

Engine designers also consider the aspect ratio of the cup-type camfollower. The longitudinal dimension of a conventional cup-type camfollower assembly must be long enough to stabilize the cam followeralong its axial length, therefore seeking to reduce the horizontalthrust exerted on the valve stem. As the cam lobe rotates and depressesthe cup, the cup's resistance to the side thrust component is manifestin wear on the cup along a line 90° from the axis of rotation of the camlobe.

In a typical cup-type cam follower, the top of the cup or cup face musthave sufficient diameter to cover the valve spring. This cupconfiguration, thus requires a cup wide enough to cover a valve springand long enough to be stable. The requirement for large cup increasesthe overall size of the assembled engine.

Crankshafts

Typical single piece and modular crankshafts have suffered harmonicbreakage problems. These problems occur when the natural frequency ofvibration of the modular crankshaft matches the frequency of impulsesapplied to the crankshaft, resulting in breakage, or can induceintolerable torsional deflections of the crankshaft.

The typical high RPM engine produces power input pulses near thefrequency range of the natural resonant frequency of the typicalcrankshaft. Thus, typical modular crankshafts tend to suffer frombreakage as the input frequency matches the natural frequency ofvibration. Typical modular and single piece crankshafts may also bedistorted and strained from bending moments asserted on the crankshaftby the force of the pistons pushing the crankshaft pins.

Cam Shafts

Typical cam shaft deflection has caused typical engine designers to haveproblems synchronizing interconnected engine modules together to achieveappropriate timing. The cam shaft twists due to the twisting torqueapplied to it, adversely affecting the timing and the synchronizationbetween engine modules. Typical engine designers utilize a large camshaft to reduce twisting of the cam shaft in an attempt to overcometiming problems. Large typical cam shaft designs, however, increases theoverall size and weight of the assembled engine.

Engine Assembly

Typical engine assembly utilizes a wide array of nuts, bolts and washersof varying shapes, sizes and lengths to assemble the parts to make atypical engine. The typical engine is assembled by different fastenerseach having different torque requirements for each individual part ofthe engine. Different fasteners and different torque create a nonuniformstress gradient on the typical assembled engine. Nonuniform stressdistorts the shape of the engine. Diversity of fasteners createsinventory overhead work for the engine manufacturer. The manufacturermust keep up with a wide variety of different size nuts and bolts. Thus,a wide variety of tools are required. Typical engines are assembledutilizing a different tool and assembly procedure for each part of theengine. Typical engines also utilize gaskets between metal parts whichcreates an assembled tolerance variation. Gaskets variably compress to anonuniform thicknesses according to the pressure applied to the gasket.The pressure varies at each fastener and at each fastener location. Thusthe tolerance of the assembled engine can vary as the thickness of thesealing gaskets vary.

When assembling modular engines designers have found that typicalengines require a different size oil pump and cooling pump for eachdifferent modular engine configuration, depending upon the number ofmodules connected to construct the engine. Oil pump size varies withengine size. Thus, the manufacturer must supply a different size coolantand lubrication pump for each configuration of one, two, three, four, orfive typical engine modules connected together to construct an engine.

Typically lubrication and coolant fluid flow serially throughinterconnected engine modules so that the lubricant and coolant fluidenter the first engine module where the fluid is pro-heated by the firstengine module before the fluid enters the second engine module, thirdmodule, fourth module, and so on. Thus, the fluid entering the lastengine module is substantially warmer than the fluid that entered thefirst engine module. Thus each typical interconnected engine modules runat a different temperature.

Pistons

Typical piston assemblies utilize a trunk style piston. The trunk pistonhas a flat circular top and a long cylindrical body or trunk. The trunkof the conventional piston fits closely within the cylinder. Thecylinder wall guides the trunk of the piston and provides for stabilityof the piston along the longitudinal axis of the cylinder. The trunk ofthe conventional piston must be long enough, relative to the diameter ofthe piston, to provide adequate stability. The ratio of the pistonlength over the piston diameter determines how stable the motion of thepiston is. The trunk of the piston rubs along the cylinder wall. Thecylinder wall guides the piston. The additional weight of the elongatedpiston trunk increases the dynamic weight of the piston, therebyincreasing the accelerative forces exerted on the piston, connecting rodand crankshaft pin.

Typical pistons such as the trunk type piston, increase the overall sizeof the engine because the length of the cylinder must accommodate theadditional length of the conventional piston trunk plus the displacementof the connecting rods. The typical trunk type piston also suffers fromthermal expansion problems. Metal expands when heated. The trunk typepiston swells to a large diameter when heated. Thus, the cylinder mustbe large enough to allow passage of the enlarged heated piston. Thecylinder diameter must be large enough to maintain a substantialclearance between the cylinder wall and the piston trunk over the fullrange of engine operating temperatures. The clearance between theoutside diameter of the conventional trunk type piston and the internaldiameter of the cylinder wall must be maintained at all operatingtemperatures or the piston will "seize up" in the cylinder. Thus,typically, a substantial gap exists between the piston trunk and thecylinder wall to allow for variations in the diameter of the piston overthe full operating temperature range of the engine. This excess gap leftto allow for swelling of the piston creates a problem. At lowertemperatures, there is a large gap between the piston trunk and thecylinder wall. At higher temperatures, the gap-between the piston andthe piston wall is very narrow. The gap between the cylinder wall andthe piston trunk, varies widely over the operating range of the engine.Thus there is a variation in the stability of the piston along thelongitudinal axis of the cylinder.

These thermal expansion considerations require engine manufacturers todesign within close tolerances yet leave large gaps to account for widevariations in piston size over the operating temperature range. Pistonstability along the longitudinal axis of the cylinder varies widely overthe operating temperature range. Moreover, high tolerance requirementsslow down the manufacturing process, to insure that the high toleranceis maintained. Slower manufacturing, requires additional man hours andtime to produce the engine.

Connecting Rods

Typically connecting rods encircle and rotate around a crankshaft pin.The connecting rod end which attaches to the crankshaft pin must be acertain minimum width so that adequate lubrication can be establishedbetween the connecting rod end and the rotating crankshaft pin.Lubrication is in adequate below this minimum width causing increasedwear and mechanical failure.

Typically engines utilize connecting rods which are open at one end andbolted to a semi circular connecting rod bracket to form a circle arounda crankshaft pin. The two piece, nut and bolt connecting rodconfiguration requires considerable additional mass for the nuts andbolts, thereby increasing the dynamic weight and forces experienced bythe crankshaft and connecting rod attached thereto.

The typical connecting rod requires considerable space. Although someengines attach more than one connecting rod to each crankshaft pin,typically the rods are side by side on a single crank pin. In thisconfiguration, each connecting rod applies a sheer force across theentire crank pin length, a distance equal to twice the width of theconnecting rod at the crank pin. The sheer force and attendant bendingmoment can cause bending and even breaking of the crankshaft pin.

Cylinder Head Seal

Some typical engines utilize a single piece head and cylinder assemblycomprising a one-piece cylinder and cylinder head. This one-piececonfiguration presents a problem in machining the cylinder head. Machinebits must extend through the length of the cylinder to reach the machinesurfaces of the attached cylinder head. Thus longer cutting bits must beused to reach the head. Longer bits are less rigid and thus reduce theaccuracy of the head machining process.

Other engines utilize a separate cylinder and cylinder head. Engineassemblers seal the cylinder head to the cylinder formed in an engineblock with large bolts and gaskets. Gaskets are subject to variablethickness, depending upon the amount of pressure applied at each boltlocation which the gasket seals. Irregular tolerances in an assembledengine decreases the structural integrity of the assembled engine. Forexample, typically, head bolt assembly methods rely on high pressures atisolated fastener points which deforms the engine block and degredatesthe structural integrity of the engine. Typical head sealing methodsrequire a complex bolt tightening pattern to exact torque requirements.Such a methodology is prone to irregular assembly.

SUMMARY OF INVENTION

In accordance with the present invention, an engine is providedcomprising one, or a plurality piston cylinders. A larger engine can beconstructed from a plurality of the engines by interconnecting engines.Interconnected engine modules are sealed utilizing an O-ring. The engineprovided by the present invention may be assembled and interconnectedwith a plurality of engines utilizing a single size uniform fastener.

In accordance with the present invention, a modular crankshaft isprovided having a crank pin comprising male and female portions. Themale and female portions interconnect to form a crank pin. Theconnections also link crankshaft sections together. The male and femalesections are splined together.

In accordance with the present invention, a piston is providedcomprising a piston having a crown. A rail guide assembly is attached tothe bottom of the piston crown. The piston rail guide assembly rides onguides formed on the engine block in which the piston resides. Thepiston rail guide assembly stabilizes the piston crown so that thepiston crown face remains perpendicular to the longitudinal axis of thecylinder in which it reciprocates. The piston is substantially smallerthan the cylinder in which it resides which reduces wear on the cylinderwall. The piston crown center is guided along the center of the cylinderby thrust pads. Thrust pads attached to the bottom of the piston crownslide along the cylinder wall to guide the center of the piston crownwithin the center of the cylinder.

In accordance with the present invention, a connecting rod is providedwhich at one end fits around a crankshaft pin and at the other endattaches to the bottom of the piston crown. The connecting rod does notfully encircle the crankshaft pin so that a plurality of connecting rodsare held in place by a retaining ring which encircles a singlecrankshaft pin within the width of a single connecting rod. Connecting aplurality of connecting rods within the width of a single connecting rodon a single crankshaft pin, shortens the overall length of thecrankshaft. A shorter crankshaft suffers less distortion duringoperation.

The other end or small end of the connecting rod is rotatably attachedto the bottom of the piston crown. The piston crown rail guide fits overand retains the small end of the connecting rod and a connecting pin.The connecting rods rotate about the connecting pin which abuts thebottom surface of the piston crown. The rail guide fits over and retainsthe connecting rod and pin under the piston crown. The connecting rodassembly shortens the overall dimensions of an engine and reduces wearon the connecting pin.

In accordance with the present invention, a cam shaft is provided. Thepresent invention provides a quill shaft which synchronizes the timingof separate and independent cam shafts which are provided in each of theseparate but interconnected engines.

In accordance with the present invention, a lubrication and coolingsystem is provided within each engine. Thus, a series of interconnectedengines are inherently equipped with an appropriate lubrication andcooling system. In accordance with the present invention a valve head isprovided which fits into the halves of an engine module. These and otherprovisions of the present invention are illustrated in the followingdescription.

The engine of the present invention provides a plurality which maybeduplicated to provide identical compact engines which may beinterconnected to form a larger engine. Each engine contains either one,two, three or more cylinders. An eight cylinder engine can beconstructed by interconnecting two four-cylinder engines or byinterconnecting four two-cylinder engines.

The engines are easily interconnected in metal to metal contactutilizing uniform fasteners and O-rings to form seals betweeninterconnected modules. The uniform fastener reduces assembly time andhelps to standardize assembly tools and methods. The modular engine usesa plurality of identical fasteners to assemble the entire engine.

The engine of the present invention provides a cam follower apparatusthat is configured to reduce the overall size of an engine while greatlyincreasing the allowable margin of error during the manufacturingprocess. Guide rails are provided on the cam follower body whichattenuate the horizontal side thrust component of the cam lobe thrust,so that the valve stem is actuated essentially by only the verticalthrust which acts parallel to the valve stem's longitudinal axis ofmotion, reducing wear.

The engine of the present invention provides a cam shaft in each enginemodule. The cam shaft in each engine module is synchronized with the camshafts in other interconnected engine modules by use of an external highRPM quill shaft. The cam shafts are geared to the high speed quill shaftwhich reduces timing errors induced by twisting of the cam shaft.

The engine of the present invention is assembled utilizing a pluralityof uniform fasteners. Using a single fastener reduces the manufacture'srequirement for inventorying of different size and length nuts andbolts. Uniform fasteners also simplify engine assembly methods. Theuniform fastener enables the present invention to utilize a large numberof uniform fasteners which evenly distribute the forces applied to theengine across the engine structure.

The engine of the present invention provides a piston crown whichutilizes thrust pads to center the piston crown within the center of acylinder. Guide rails which run within guide slots are attached to thepiston crown. These guide rails keep the piston crown face stable alongthe longitudinal axis of the cylinder. The stabilizing influence of thepiston guide rails eliminates the need for the long piston trunktypically used in engines. The piston enables an engine manufacturer toassemble an engine which is smaller than a typical engine with the samestroke. This present invention provides a structure which reduces oreliminates the bending moment of the shear force acting on theconnecting rod pin. Thus, the size and weight requirements for theconnecting rod pin is reduced. The reduced size and weight of the pinconnecting rod assembly reduces the dynamic weight and wear on the pinduring operation of the piston assembly.

The placement of the connecting rod abutting the lower surface of thepiston crown enables the connecting rod to pivot close to the pistoncrown upper face. This configuration shortens the distance between theconnecting rod end and the piston crown upper surface, which provides anengine smaller than a typical engine with the same stroke. Thrust padsare utilized to maintain the piston crown within the center of thecylinder.

In the engine of the present invention, a plurality of connecting rodends are connected to a crankshaft pin. The connecting rod end hassubstantially the same diameter and radius curvature as the crankshaftpin. A circular bearing between the crankshaft pin and the connectingrod end facilitates lubrication. A set of retaining rings is provided tomaintain contact between the crankshaft pin and the connecting rod endassembly.

Attaching more than one connecting rod end to a single crankshaft pinreduces the overall length of the crankshaft, which reduces the bendingmoment of the shear forces applied to the crankshaft by the pistonsthrough the connecting rods. Reducing the bending moments induced in thecrankshaft pins, by reducing their length overall, increases thestructural integrity of the crankshaft during operation. The crankshaftis shorter than a typical crankshaft.

The present invention provides a large diameter crank pins andcrankshaft to reduce twisting and torsional deflections induced in thecrankshaft. A tab on the connecting rod bearing restricts the rotationalmotion of the bearing relative to the connecting rod ends so that oilsupply apertures in the tabbed bearing are not exposed to the gapsbetween the connecting rod ends.

Cylinder Head seal

The cylinder head of the present invention is configured to facilitatemachining of the intake ports, exhaust ports and valve guides in thecylinder head. The top of the cylinder is cut at an angle so that theline at the top edge of the angled cylinder edge creates a high loadingwhen pressure is applied. This enables the angled cylinder edge to forma metal to metal seal against the cylinder head. A circumferential landaround the cylinder, circumferential land around the cylinder head and aretaining band are utilized to attach and seal the cylinder to thecylinder head the retaining ring and lands fit into a receiving grovecut in each engine block half.

Lubrication system

The engine of the present invention provides an independent lubricationsystem for each engine. Each engine contains its own independentlubrication and cooling system comprising a coolant pump, a scavengerpump, and a pressure pump. The oil supply is manifolded in parallel toeach engine so that each engine is supplied with oil at the sametemperature. Each engine module runs at the same temperature. Aplurality of modules connected together to form an extended modularengine will have an appropriate lubrication system. A main supply linefrom the oil radiator outlet is manifolded in parallel through aconstant temperature line into each of the engines so that thetemperature of the oil at each engine is the same.

Crankshaft

The crankshaft is comprised of a plurality of modules which interconnectin a male-female fashion to form a crankshaft. The male-femalecrankshaft connections are splined together for rotational stability.The crankshaft is made of a stiff material with a large diameter so thatthe natural frequency of vibration of the crankshaft is much higher thanthe frequency of the rotational impulses applied to the crankshaft bythe low RPM engine. Thus, the frequency of piston impulses does notenter the range of the crankshaft's natural frequency of vibration. Thissubstantially reduces the probability of harmonic breakage problems dueto piston impulses matching the natural frequency of vibration in acrankshaft.

Valves

The cylinder head of the present invention uses three intake and threeexhaust valves for each cylinder.

Other advantages and features of the invention will be apparent afterstudying the following description of a preferred embodiment.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross section of a single piston engine module embodying theinvention.

FIG. 2 is a cross section of a dual piston engine module embodying theinvention.

FIG. 3 is a cross section of a three piston engine module embodying thepresent invention.

FIG. 4 is a simplified plan view of four engine modules connectedtogether according to the present invention.

FIG. 5 is an exploded view of an engine module embodying the invention.

FIG. 6 is a simplified schematic of a multi-cylinder intermeshed set ofengine modules showing pistons in dotted lines.

FIG. 7A is a perspective of a cam follower embodying the presentinvention.

FIG. 7B is an elevation of the cam follower shown in FIG. 7A in contactwith a cam shaft lobe.

FIG. 7C is an elevation of a cam lobe and follower according to theinvention.

FIG. 7D is a plan view in direction 7D illustrated in FIGS. 7A and 7B.

FIG. 7E is a force vector diagram.

FIG. 8 is a cross-section of a prior art cup-type cam follower assembly.

FIG. 9 is a plan view of a cam follower guide rail.

FIGS. 10A is a plan view of a modular cam shaft embodying the invention.

FIG. 10B is a section along line 10B--10B in FIG. 10A.

FIG. 11 is a perspective view of dual cam shafts on a simplified enginehead.

FIG. 12 is a plan view of several cam shafts connected together using aquill shaft.

FIG. 13A is an exploded view of a piston that embodies the invention.

FIG. 13B is a section along line 13B--13B in FIG. 13A.

FIG. 13C is a section taken along line 13C--13C in FIG. 13B.

FIG. 14 is a section taken along line 13C--13C in FIG. 13B.

FIG. 15A is a plan view of a crank pin in a conventional prior artengine.

FIG. 15B is a plan view of a crank pin in a conventional prior artengine.

FIGS. 16A and 16B are enlarged plan views, 90° apart, of the pistonshown in FIG. 13A.

FIG. 17 is a plan view of a connecting rod assembly embodying theinvention.

FIG. 18A is a plan view of an assembled connecting rod assembly andcrank pin.

FIG. 18B is a sectional taken along line 18B--18B in FIG. 18A.

FIGS. 19A and 19B show relative positions of connecting rods and atabbed bearing at two different times in an engine power cycle havingtwo pistons.

FIG. 20 is a plan view of two disassembled sections of the male andfemale crank pin sections of a crankshaft according to the invention.

FIG. 21 is a plan view showing a connecting rod on the crankshaft ofFIG. 20.

FIG. 22 is a sectional view of a portion of a cylinder head using a sealarrangement according to the invention.

FIGS. 23A and 23B are sectionals of the cylinder head shown in FIG. 22with a retaining ring in different stages of deformation.

FIG. 24 is a plan view of a portion of a cylinder head.

FIG. 25 is an elevation of an exhaust valve and gas port embodying thepresent invention.

FIG. 26 is an elevation of the exhaust valve and gas port also showing avalve retraction mechanism and a valve in a cylinder head.

FIG. 27 is an elevation of one-half of an engine block according to theinvention.

DESCRIPTION OF AN EXAMPLE OF A PREFERRED EMBODIMENT OF THE INVENTION

Turning now to FIG. 1, in the present example of a preferred embodimentof the invention, the engine of the present invention provides aplurality of identical engines which may be interconnected to form alarger engine. As shown in FIG. 1, each engine contains either one 10,two 12, three 14 cylinders or more. Each cylinder houses a piston, e.g.16, 18, 20. As shown in FIG. 4, these individual engines may beinterconnected by abutting the planer surfaces 25 located mid-way 24between the axial separation of the cylinders in adjacent modules. Eachengine provides one piston 16, two pistons 18, or three pistons 20. Theengines run independently or may be interconnected to work incooperation.

An eight cylinder engine can be constructed by interconnecting fourtwo-cylinder modules 12 or eight one-cylinder modules 10. As shown inFIG. 4, the engines are interconnected utilizing metal to metal contactat the axial plane 24 mid-way between adjacent cylinders 22. A metal tometal contact is formed between the adjacent planer surfaces 26utilizing uniform fasteners discussed below. An O-ring groove isfashioned in the planer surface 26 of each engine. The O-ring is placedin the O-ring groove to form an O-ring seal between adjacentinterconnected engines.

As shown in FIG. 5, each engine is split into two halves 26 on a planeperpendicular to the longitudinal axis of the crankshaft 30. A groove195 is formed on the interior wall 147. A cylinder head 138 and cylinder141 are fastened together with a retaining ring 142 that fits into thegroove 195 and is secured when the two halves 26 are brought together(see FIGS. 22, 23A, 23B and 24). In the present example of a preferredembodiment, the engine utilizes a piston crown and piston guide railassembly, rather than a trunk type piston. The piston crown assemblyenables a designer to reduce the size of the engine and prolongs enginelife by reducing induced wear. The piston crown assembly is stabilizedby guide rails and thrust pads instead of the piston trunk.

Referring to FIG. 6, the cylinder spacing within an engine is configuredso that an engine can be intermeshed with an adjacent engine. The splitplane of one engine becomes the separation plane between theintermeshing engines. Single cylinder engines can be interlaced into 2cylinders, 4 cylinders, 6 cylinders, etc., configurations. Threecylinder configure engines can be similarly interlaced as 3 cylinders, 6cylinders, 9 cylinders, 12 cylinders, etc., engine configurations. Inthe case of the meshed configuration engine, an extra crank throw isintroduced between bearings. All other interfaces remain identical,differing only in axial dimension.

In the present example of a preferred embodiment of the presentinvention, each engine provides a lubrication and coolant system and acam follower apparatus. When a plurality of engines are interconnected,it becomes desirable to synchronize the firing of the pistons in theindividual engines. Synchronizing enables proper timing of the overallcomposite engine composed of a plurality of engines running insynchronization. Therefore, the individual cam shafts in each engine aresynchronized. In the present example of a preferred embodiment of thepresent invention, synchronization between the plurality of enginesinterconnected is facilitated by an external high RPM quill shaft,discussed below.

In the present example of a preferred embodiment of the presentinvention, the engine utilizes a guide slot to stabilize the piston andguide rails to stabilize the cam follower mechanisms along theirrespective axis of translation. The guide slots and rail guides of thepresent invention are compact and require less space to perform theirrespective function than typical equivalents. Compact design for theguide slots and rail guides reduce the overall size of the engine andprolong its useful life.

The entire engine can be assembled and interconnected with other modulesutilizing a single uniform fastener and tool, discussed below. In thepresent example a preferred embodiment of the present invention, theengine utilizes a large diameter modular crankshaft, discussed below. Ina preferred embodiment of the present invention, the modular engine isassembled utilizing a uniform fastener of constant size and length. Thefastener positions 200 present in an engine are illustrated in FIGS. 27.The engines are connected in metal to metal contact providing a uniformcumulative assembled tolerance for the final assembled engine. Uniformcumulative assembled tolerance enables an engine manufacturer tointerconnect a plurality of engines without experiencing cumulativetolerance errors between the engines. Cumulative tolerance errors may beexperienced when a series of engines are interconnected with gasketswhose thickness may vary according to the force applied. The cumulativeerror experienced when gaskets are used, may become significant wheninterconnecting a stack of engines such eight two-cylinder engines,which could be interconnected to form a sixteen cylinder (V-16) engine.Cumulative tolerance errors may cause the engines to align improperlywith the crankshaft, due to a variation in the longitudinal axis of thecrankshaft. The metal to metal contacts of the present invention enablethe eight engine stack for example, a V-16 to be uniform along thelongitudinal axis of the crankshaft, without variations caused by thecumulative tolerance errors which may be caused by assembling withgaskets.

In the present example of a preferred embodiment of the presentinvention, the engine utilizes the entire facial cross section of anengine to form a metal to metal contact, and O-ring form a seal betweenthe entire facial cross sections of adjacent engines. Unlike theVoorhies modular engine discussed earlier, the engine of the presentexample of a preferred embodiment provides for metal to metal contactbetween entire cross sections of adjacent engines, enabling the presentinvention to achieve a more compact design along the longitudinal axis,that is, build a shorter engine.

Cam Follower

Turning now to FIG. 7, in the present example a preferred embodiment ofthe present invention a cam follower 32 is utilized to reduce theoverall size of the engine and increase its useful life. As shown inFIG. 7B, cam follower mechanism 32 acts as a mechanical intermediarybetween the rotating cam shaft lobe 40 and the valve stem 54. As shownin FIG. 7C, cam shaft lobe 40 rotates about cam shaft axis 46. As shownin FIG. 7D, the guide rails 34 of cam follower 32 slide up and down inguide slots 37. Guide rails 34 are formed on the sides of the camfollower 32. Guide slots 37 are formed in the cylinder head 35 and acylinder filler block 33, which is installed or formed in the cylinderhead.

As shown in FIG. 7E, the thrust from the rotating cam lobe 40 may beresolved into a horizonal component 39 (side thrust) and a verticalcomponent 41 (down thrust). The cam follower guide rails 34 absorb theside thrust component 39. Thus, only the down thrust cam lobe thrustcomponent 41 is transmitted through the cam follower mechanism 32 to thevalve stem 54. Reduction of side thrust reduces wear on a valve stem,for example, valve stem 54 and any associated valve guide.

Guide rails 34 are utilized in the present example of a preferredembodiment to absorb the side thrust component 39 and to providestabilization of the cam follower 32 along the axis of translation. Thecam follower slides up and down on an axis parallel to the longitudinalaxis of the cam follower guide rails 34.

The cam follower reduces wear on the valve stem by attenuating the sidethrust component 39 of the cam lobe thrust. Thus, only vertical thrust,parallel to the longitudinal axis of the valve stem, is asserted on thevalve stem reducing wear thereon. Side thrust increases wear on thevalve stem and thus reduces engine life. The cam follower mechanism 32of the present invention operates in an oil lubricated environmentwithin the cylinder head.

Unlike typical cup-type cam follower mechanisms, as shown in FIG. 8, thepresent invention relies on the aspect ratio of the cam follower guiderail 34 rather than the aspect ratio of the diameter of the conventionalcup-type cam follower 48. The cup-type cam follower 48 relies on itscup-shape for stability. The cup acts as a mechanical intermediarybetween the cam lobe 40 and a valve spring 50. Cam lobe 40 rotates aboutcam shaft axis 46. Cam lobe 40 depresses cup-type cam follower 48, whichin turn depresses valve stem 54. Valve stem 54 is depressed along thelongitudinal axis of the valve stem 54, and guided by valve guide 52.The conventional cup-type cam follower 48 relies on the aspect ratiodefined by the diameter of the cup over the length of the cup, toachieve stability of the cam follower along the longitudinal axis oftranslation of the valve stem 54. The diameter 43 of the cup-type camfollower 48 must be large enough so that it will fit over the valvespring 50, or some other valve return mechanism. Therefore, the minimumdiameter 43 of a cup-type cam follower must be slightly larger than thediameter of the valve spring 50. Thus, the diameter of the valve springdictates the minimum length of side 49 of the cup required to stabilizethe cup. The large minimum diameter cup dictates a long minimum cuplength, which increases the size of the engine. Typical engine designsutilize long valve stems to increase the aspect ratio of the valve stemand reduce engine wear. Long stems increase the overall size of theengine. The engine of the present invention provides compact short stemvalve and valve stem.

Referring back now to FIG. 7A, in the present example of a preferredembodiment of the present invention, the engine provides cam follower32. Cam follower 32 relies on the aspect ratio of the guide rail 34 toabsorb the side thrust and to achieve stability along the longitudinalaxis of the valve stem. Cam follower 32 of the present invention doesnot have to fit over the valve spring as does the typical cup-type camfollower. This enables the cam follower of the present invention toprovide a compact cam follower which reduces the required size of thecam follower and thus reduces the overall size of the engine in which itis embodied.

Cam follower 32 of the present invention relies on the aspect ratio ofcam follower rail 34 (the ratio of the length divided by the width ofcam follower rail 34) for stability. Cam follower guide rail width issignificantly less than that required in a cup-type cam follower, whichmust fit over the valve spring. The cam follower guide rail of thepresent invention dues not have to fit over the valve spring andtherefore is much smaller. Because the width of the cam follower guiderail 34 is significantly less than the required diameter of the cup-typecam follower, the cam follower of the present invention enablesconstruction of a structure which provides high aspect ratio for the camfollower guide rail, yet utilizes significantly less space for any givenaspect ratio.

To maximize the guide rail aspect ratio, and the stability of the guiderail 34 the end 42 of the guild rail 34, which engages the cam shaft 43,as shown in FIG. 7B, is cut out to match the diameter of the cam shaft43, which it engages. This maximizes the length of the face of guiderail 34 adjacent guide rail slot 37. The longer rail length absorbs moreside thrust and provides more stability to the cam follower along thecam follower's axis of translation. Thus, the cam follower is small buteffectively attenuates the side thrust of the cam lobe. FIG. 9 is adetailed illustration of the cam follower guide rail 34, interface 42,and the cam shaft 43.

In the present example, the modular engine of the present invention runsat approximately 2700 RPM. The cam shaft RPM is approximately 1350. Thelower RPM and compact design cam shaft reduces the accelerative forcesasserted on the cam shaft, the cam follower and the valve assembly.Thus, the cam shaft can be easily constructed by pressing cam lobes ontothe cam shaft to obtain an elastic fit, rather than using typical slowermanufacturing methods which utilize a plastic fit. The reducedaccelerative forces enable the engine to provide a compact and lowpressure valve/valve spring apparatus. Thus, the engine provides asmaller diameter valve face, and a shorter length valve stem thantypical valves. This compact design valve substantially reduces thedynamic mass of the valve of the present invention over that of typicalprior art valve assemblies.

Typical valves are long in order to enhance the stability along its axisof translation. The cam follower of the present invention efficientlyabsorbs the side thrust component of the cam lobe thrust so that lesslongitudinal stability compensation is required by the valve stem. Thus,the valve stems do not have to be as long because they do not have tocompensate for instability, as are required by the typical valve stems.Thus the present invention valve reduces the required overall size ofthe engine.

In the present example of a preferred embodiment, the engine utilizessix valves per engine. Utilizing six valves and a low RPM creates a verylight valve requirement and with low inertia. Cam lobes can thus bestamped from sheet metal or made as powered metal pressings and pressedonto the cam shafts as shown in FIGS. 10A and 10B.

Quill Shaft

Turning now to FIG. 11, in the present example of a preferredembodiment, each engine 10 provides two cam shafts 56 and 58. Each camshaft provides three lobes 60. The rotation of cam shafts 56 and 58 issynchronized by gear 62. FIG. 12 illustrates a series of interconnectedengines 10. The timing of the cam shafts 56, 58 for each module issynchronized to the timing of the cam shafts in other interconnectedmodules.

As shown in FIG. 12, in the present example of a preferred embodiment,the present invention utilizes a quill shaft 64 to synchronize theplurality of cam shafts 56 and 58 provided by each interconnectedengine. The quill shaft is driven by step-up drive 66, which is attachedto and driven by the crankshaft 68. The step-up drive 66 enables thequill shaft 64 to run at substantially higher RPM than the crankshaft.

In the present example of the preferred embodiment of the presentinvention, the quill shaft RPM is twelve times that of the crankshaft.Quill shaft 64 comprises a plurality of interconnected sections 65. Eachindividual quill shaft section 65 is coupled to an individual engine camshaft. The high RPM quill shaft reduces the torque for a given appliedforce exerted on the quill shaft. The torque exerted on the quill shaftis reduced by a factor of twelve or the ratio of the quill shaft RPMdivided by the crankshaft RPM. The reduced torque induces less torsionaldeflection or twisting for a given horse power input, than it would at alower RPM and the same applied horse power. Thus timing errors beinginduced by torsional deflections are significantly reduced or eliminatedby the reduced torque, high RPM quill shaft.

The quill shaft 64 of the present invention enables selection of avariable quill shaft size to accommodate a specified tolerable torsionaldeflection, or timing error, for an engine comprised of a given numberof interconnected engines. Each individual engine is identical, thuseach engine provides the same valves, crankshafts, cam shafts and camlobes. The external quill shaft enables the engine designer to useidentical engines to build up larger engines, and maintain independentcontrol over timing errors between the engines by introducing a quillshaft to synchronize the timing between the engines.

Piston

The piston of the present invention enables the manufacture to assemblea engine which is smaller than a typical engine having the same stroke.Because the connecting rod is attached near the piston face at the lowersurface of the piston crown, the engine cylinder length need accommodateonly the stroke or axial displacement of the piston, without providingthe additional length necessary to accommodate the trunk of a typicalpiston. The preferred embodiment of the piston assembly provides asmaller connecting pin than a typical piston. The engine enables asmaller pin to be utilized by reducing stress forces on it. The smallerpin reduces the dynamic weight of the piston assembly and, theassociated accelerative forces asserted on it, thus reduces theconnecting rod, and the crankshaft to which it attaches.

Placement of the connecting rod abutting the lower surface of the pistoncrown enables the connecting rod to be attached close to the upper faceof the piston crown, thereby shortening the distance between theconnecting rod end and the piston crown. In the typical engine design,the distance between the piston face and the connecting rod end isincreased by the length of the piston trunk. Thus, the piston of thepresent invention does not require as much space to accommodate the sameengine stroke because the present invention does not have to accommodatethe additional length of the piston cylinder trunk.

The piston crown of the present invention seals the combustion chamberutilizing a piston ring. The piston crown does not rub against thecylinder walls. The piston crown utilizes thrust pads to slide along thecylinder wall guiding the center of the piston. The piston crown can bemade of a material which expands and contracts readily under the varyingtemperatures experienced during engine operation. When the engine firststarts, it is cold and the gap between the cylinder wall and the pistonis relatively large. The piston crown contracts and expands. The crownis made of thermal conductive material which disburses heat withoutconcerning the engine designer with the clearance between the pistoncrown edge and the cylinder wall.

Turning now to FIGS. 13A-13C, the stability of the piston crown face asperpendicular to the longitudinal axis of the cylinder is provided bythe cross head rail guide assembly, rather than the typical pistontrunk. The present invention provides guide rails 84 and keys 82 tostabilize the piston crown face perpendicular to the longitudinal axisof the cylinder as shown in FIGS. 13A, 13B, 13C and 14.

The stability of the piston face is dependant upon the aspect ratio ofthe stabilizing member. Typically, the piston trunk must be long enoughrelative to the diameter of the piston face to obtain a suitable aspectratio and associated stability of the piston face with respect to thelongitudinal axis of the cylinder. The present invention utilizes across head guide rail assembly to provide stability to the piston crownface in a plane perpendicular to the longitudinal axis of the cylinder.Thus, it is the dimensions of the small rail guide rather than thelarger piston diameter which dictate the aspect ratio and stability ofthe piston crown in the present invention. The present inventionprovides greater stability utilizing a smaller space, because thestability of the piston crown in the present invention depends on thedimensions of the guide rails 84 rather than the dimensions of thepiston.

The stability of a typical piston varies over the operating range of theengine, because the clearance between the stabilizing member, the pistontrunk, and the cylinder wall, varies as the piston expands and contractsunder temperature variations. The typical engine designer must allowsufficient space between the piston trunk and the cylinder wall toaccommodate the expanded piston when hot and swollen. At coolertemperatures, the cooler piston has a smaller diameter. There is alarger gap between the cylinder wall and the piston trunk. Thus, thereis less stability of the piston at lower temperatures when the pistoncools than when it is hot. The stability of the typical piston variesover the operating temperature, as the gap between the piston trunk andthe cylinder wall varies, when the piston expands and contracts.

The rail guide assembly of the present invention maintains a much moreconsistent stability over varying temperatures. The rail guide is lesssensitive to temperature variations. The rail guide relies on thesmaller dimensions of the guide rail and associated guide key tomaintain tolerances over a wide thermal and stable temperature of thepiston. The width of the guide rail, utilized in the present invention,is substantially smaller than the width of a trunk type piston. Theguide rail assembly of the present invention is less affected bytemperature ranges because there is less metal to expand.

In a preferred embodiment of the present invention, the guide rail isone twenty-forth (1/24th) as wide as a trunk type piston. Thus, theguide rail will expand twenty-four (24) times less than a trunk typepiston having a diameter twenty-four times as wide as the guide rail andmade of the same material, operating at the same temperature. Thus, anexcess gap between the stabilizing member (the guide key) and the guiderail in which it resides could be twenty-four (24) times smaller thanthe gap between the conventional stabilizing member (the piston trunk)and the guiding cylinder wall in which it resides. This factor oftwenty-four (24) gap tolerance clearance advantage manifests itself inthe tolerance of the manufacturing process. The engine, of the presentinvention can be manufactured using reduced tolerance machining toenable manufacture of the engine to proceed quickly and without excesstolerance or induce wear caused by a loose fitting piston rattling in atypical engine. The piston of the present invention maintains a moreconsistent stability and decreases engine wear while enabling an overallsmaller engine to be assembled.

Turning now to FIG. 13A, the engine provides a piston crown 70 and crosshead guide rail assembly 72. Piston thrust pads 74 are provided tocenter the crown in the cylinder. Connecting rod 76 engages connectingrod pin 78, which abuts bottom of piston crown 70. Cross head rail guide72 is attached to piston crown 70 by bolts 80. As shown in FIG. 13C,guide keys 82, are provided in the lower crank case where guide keys 82engage cross head guide rail slots 72. The cross head guide rail slots72 and keys 82 stabilize piston crown face 86 and keep it perpendicularto the longitudinal axis of the cylinder parallel to the axis oftranslation of the piston crown within the cylinder. A plane drawn inthe face of the piston crown 86, thus is kept perpendicular to thelongitudinal axis of the cylinder. Thrust pads 74 maintain the pistoncrown 86 centered within the cylinder in which it resides. FIG. 13C is asectional view of the rail guides 72 and keys 82. The slot guide passesalong side the crankshaft enabling a shorter height cylinder and thus ashorter engine to be manufactured using the engine of the presentinvention.

FIG. 14 shows the cross head guide rail assembly for the piston. Theguide slot 72 of the cross head guide assembly engages the key 82. Asshown in FIG. 14, the guide rail keys 82 protrude along the crank casewall along the longitudinal axis of the crankshaft 68. The center line88 of the cylinder is shown for reference.

The piston utilized in the present example of a preferred embodimentprovides several advantages over typical trunk type pistons. As shown inFIGS. 15A and 15B, the typical connecting rod pin 96 is located adistance 98 below the bottom of the piston 94. Typical connecting rods100 are attached to connecting pin 96. In the typical trunk type piston,as shown in FIG. 15A, the force of combustion 92 presses down on the topof conventional trunk type piston 94. As shown in FIG. 15B, thecombustive force 92 pressing down on conventional piston top 94 places aforce and induces an associated bending moment on connecting pin 96.This bending moment tends to stress connecting pin 96, trying to bendconnecting pin 96 around the longitudinal axis 102 of connecting rod100. This bending moment tends to place undue wear on the connecting pinand shortens engine life. There is no bending induced on the connectingpin of the piston assembly provided by the present invention.

Turning now to FIG. 16A, in the piston of the present invention, force92 acts on the top of the piston crown 86. In the present inventionconnecting pin 78 adjoins both the lower surface of the piston crown 86and the top of connecting rod 76. Thus, there is no bending momentapplied to the connecting pin 78, as it mechanically engages both thebottom of the piston crown 86 and connecting rod 76. The connecting rod76 and connecting rod pin 78 are attached to piston crown 86 byretaining rings 104. FIG. 16B is a view of the piston crown connected tothe connecting rod and connecting pin turned ninety degrees (90° ) fromthe view shown in FIG. 16A.

Connecting Rod

Turning now to FIG. 17, in the present example of a preferred embodimentof the present invention, the engine provides a connecting rod 76. Thesmaller end 116 of connecting rod attaches to the bottom surface of thepiston crown as shown in FIG. 13A, discussed earlier. The large end ofthe connecting rod 107, as shown in FIG. 17, connects to the crank pin108, as shown in FIGS. 18A and 18B.

Turning now to FIGS. 19A and 19B, in the present example of a preferredembodiment, the large end of the connecting rod forms a 136°semicircular arc which closely approximates the outside diameter oftabbed bearing 124. Tabbed bearing 124 abuts connecting rod end 107 onits outside diameter and the crank pin 108 on its inside diameter. Thetabbed bearing 124 provides oil apertures 122 which enable oil to passto provide lubrication for the connecting rod crank pin assembly.

As shown in FIG. 18B, the width 126 of connecting rod end 107 ispreferably a minimum distance to enable hydrodynamic bearings to beformed between the connecting rod end 107 and the tabbed bearing 124.The tabbed bearing also provides for lubrication between the internaldiameter of the tabbed bearing 124 and the crankshaft pin 108.

As shown in FIGS. 19A and 19B in the present example of a preferredembodiment, a plurality of connecting rods 76 are attached to onecrankshaft pin 108. Connecting rods 107 preferably do not encirclecrankshaft pin 108. Thus each connecting rod end 107 requires less than180° of crank pin surface. In the present example, they are each at a136° arc. The connecting rod ends 107 may rotate relative to crankshaftpin 108 without interfering with each other. A set of retaining rings(not illustrated for clarity) are utilized to rotationally attachconnecting rods ends 107 around the tabbed bearing 124 and thecrankshaft pin 108. The modular crankshaft pin comprises a male andfemale member which are inserted through the circular opening in thetabbed bearing 124 after the connecting rods 76 and retaining rings(FIG. 18A) have been assembled to form a circular structure around theconnecting rod assembly.

The connecting rod assembly of the present invention enables an enginedesigner to connect more than one connecting rod to a single crank pin.Multiple connecting rods can be attached to a single pin while utilizinga minimum length crank pin just long enough to accommodate lubricating aconnecting rod of minimum width 126, as shown in FIG. 18B. The minimumcrank pin length is preferably equal to the minimum width for which asingle connecting rod 107 has adequate lubrication. The minimum widthcrank pin of the connecting rod assembly enables the engine designer tobuild a shorter crank pin and overall shorter crankshaft. Eachcrankshaft pin length in the crankshaft is reduced by a factor equal tothe number of connecting rod ends attached to an individual crankshaftpin. A shortened crank pin reduces the bending moment asserted on thecrank pin. The shorter crankshaft experiences smaller bending momentsfor a given force than a longer crankshaft.

Turning now to FIG. 20, the crankshaft pin 10S is provided having largediameter crankshaft pins 109 and crankshaft 114 to reduce the torsionaldeflection induced in the crankshaft by the forces applied by thepistons.

The crankshaft provided is a plurality of modules which plug together.After the connecting rod assembly has been assembled, male and femalesections of the crank pin can be inserted and Joined inside of thecircular end of the connecting rod assembly. The connecting rod ends 107do not fully encircle the crankshaft pin 108 so that gaps 110, 112 and114, as shown in FIG. 18A, are formed between the connecting rod ends107. Tabbed bearing 124 is utilized for lubrication between thecrankshaft pin 108 and the connecting rod end 107. Tab 120, on tabbedbearing 124, restricts the rotational motion of the tabbed bearing 124and oil apertures 122 relative to the connecting rod ends 107. Thus, oilapertures 122, which supply oil to the exterior surface of the tabbedbearing and the interior surfaces of the connecting rod end 107 areprevented from rotating far enough to become exposed to the gaps 110,112, 114 between the connecting rod ends 107. Thus, oil is preventedfrom being pumped from the oil apertures 122 and through gaps 110, 112and 114. Oil pumped through the gap flows to the bottom of the engineand has to be recovered with a scavenging pump. Reduction of the amountof oil escaping through the gaps reduces the amount of oil that has tobe pumped back. The scavenging pump can be smaller in the presentinvention. This reduces the over all engine size. The positioning of theoil apertures 122, so that they stay under the connecting rod ends 107,and do not allow oil to escape through the gaps 110, 112 and 114. Thisalso reduces the amount of oil which must be supplied to the connectingrod assembly by the supply pump. This reduces the size of the oil supplypump required to pump oil to the connecting rods and thereby reduces theoverall size of the engine.

In an alternative embodiment of the present invention, a male eyelet andfemale circular eyelets are formed in the bottom of connecting rodswhich share a crankshaft pin. The female circular eyelet comprises aforked set of circular eyelets which slide over the male circulareyelet. The combined male and female eyelets form a circular eyeletconnecting rod assembly. The assembly is of a width sufficient to enableformation of hydrodynamic bearing between the crank pin and theconnecting rod ends that slides over the crankshaft pin.

In the alternative embodiment, a pressed sleeve bearing is press-fittedonto the forked female eyelet so that the pressed bearing sleeve doesnot rotate relative to the female connecting rod. The crankshaftconnecting pin rotates underneath the sleeve bearing of the female rod.The displacement between the male connecting rod end and the sleevebearing pressed into the female connecting rod end is thus minor. Themale rod rotates over the bearing fixed in the female eyelet as thecrankshaft rotates in a circle within the bearing fixed in the femaleeyelet. Two connecting rods drive two pistons by driving a singleconnection formed by the male and female connecting rod ends.

The path of the connecting rod ends 107 and tabbed bearing 124 utilizedin the present example of a preferred embodiment is illustrated in FIG.19A and 19B. FIG. 19A shows the connecting rod ends 107 and tabbedbearing 124 when the crankshaft has rotated to bottom of the pistonstroke. At this point, the center of the crankshaft is at point 130, asshown in FIG. 192A. In FIG. 19B, the crankshaft has now rotated to thetop of the piston stroke and the center of the crankshaft pin is nowlocated at point 132. Point 120 is repeated for reference. Notice thatin FIG. 19A, when the pistons are at the bottom of the stroke, theconnecting rod ends 107 do not meet but leave a gap 110 between them. InFIG. 19B, when the pistons are at the top of their stroke, theconnecting rod ends 107 rotate so that they leave small gaps 114 and 112(FIG. 18A) between the connecting rod ends 107 and tab 120 of tabbedbearing 124. Oil apertures 122 remain underneath connecting rod ends 107and are not exposed to gaps 110, 112 or 114 during any point of therotation of the crankshaft pin.

Modular Crankshaft

Turning now to FIG. 20, in the present example of a preferredembodiment, the engine utilizes a modular crankshaft 114, as shown inFIG. 20. The modular crankshaft 114 utilizes a male 111 female 109assembly to form a crank pin 108. The male section 111 slides into thefemale section 109 to form crank pin 108. The male and female sectionsare splined together for rotational fixation between them. The presentinvention provides a structure which reduces the bending moment assertedon the crank pin 108 by the connecting rod end 107. This is accomplishedby reducing the width 126 (FIG. 18B) of crank pin 108, to the minimumwidth needed to form a hydrodynamic bearing, based on the width of asingle connecting rod end 107, tabbed bearing 124, and crank pin 108.The necessary length of the crank pin is reduced because more than oneconnecting rod end 107 is attached to the pin 108.

Two connecting rod ends 107 are connected to a single width 126 crankpin 108, reducing the necessary overall length of crank pins by a factorof two, because two connecting rod ends are sharing the same crank pinwhose length equals the minimum width 126 of a single crankshaft pin. Ifthree connecting rod ends 107 are connected to a single crank pin 108,the pin length requirement is reduced by a factor of three, and so on.

As shown in FIG. 21, reducing the crank pin length reduces the overallcrankshaft length and thus reduces the bending moment asserted acrossthe width of a crank pin by connecting rod 107. Reducing the bendingmoments by minimizing the width utilizing a single width crank pin formultiple connecting rod ends, reduces the length of crank pins and thusreduces the overall length of the crankshaft. The reduced width of thecrank pins reduces the bending moment of a force asserted on a crankpin. Thus, the crank pins suffer less deformation twisting, andtorsional deflections during operation. Crank pin 108 and crankshaft 114are formed of large diameter tubing which minimizes the torsionaldeflection within the crankshaft and crank pins.

The present invention provides sufficient overlap between the malesection 111 of the crank pin and the female section 109 of the crankpin. The crankshaft is made of a stiff material and is configured inlarge diameter so that the natural frequency of vibration of thecrankshaft and crank pins is much higher than the frequency ofrotational power impulses applied to the crankshaft by the low RPMpistons through the connecting rods. The present example of a preferredembodiment, utilizes a modular engine with a maximum RPM ofapproximately 2,700. Thus, the frequency of piston impulses applied tothe crankshaft is much lower in the low RPM engine than the naturalfrequency of vibration of the large diameter crankshaft. The frequencyof the impulses supplied by the pistons does not match the naturalfrequency vibration of the crankshaft of the present invention. Thismismatch substantially reduces the possibility of harmonic breakage ofthe crankshaft to lower than that encountered with typical modularcrankshafts.

As shown in FIG. 21, crankshaft bearings 134 and 136 support eachsection 115 of the modular crankshaft. Each section of the modularcrankshaft is support by bearings 134 and 136, so that the bendingmoments and shear forces from the piston are resolved in a redundantmanner by each of the crankshaft sections 114 and 115, which connecttogether to form a crank pin 108, which receives the load from a piston.

Cylinder Head Seal

Turning now to FIG. 22, in a preferred embodiment of the presentinvention, the engine utilizes a metal to metal seal between cylinderhead 138 and cylinder 141. Cylinder head 138 is configured separatelyfrom cylinder 141. The shell of configuration of the separate cylinderhead 138 enables conventional machine bits to traverse the depth of thecylinder head 138. The shallow depth of the cylinder head 138 enablesshort rigid machine bits to accurately machine the cylinder headsurfaces. Longer machine bits, which would be required with a one-piececylinder head and cylinder would have to traverse the length of thecylinder to reach and machine the cylinder head. Configuration wouldrequire long machine bits which would be less rigid and thus lessaccurate in machining of the cylinder head 138.

The cylinder head of the present invention utilizes a metal to metalseal between chamfered edge 146 of cylinder 141 and a flat surface 150within the female portion of cylinder head 138 into which the top maleportion of the cylinder inserts. The present invention has advantagesover cylinders assembled using gaskets to seal the cylinder head. Themetal to metal contacts of the present invention forms a seal withoutthe attendant variations in assembled tolerances experienced whenutilizing gaskets to assemble an engine.

The cylinder head 138 forms a female receptacle into which the wall ofcylinder 141 slides and mechanically engages. Chamfered edge 146 ofcylinder wall 141 abuts flat surface 150 of the cylinder head 138.Turning now to FIG. 23A, retaining ring 142 is shown as a U-shapedbracket, forming right angles 144 and 143, and fitting over cylinderhead land 139 and cylinder wall land 140. Retaining ring ends 143 and144 abut lands 139 and 140.

Turning now to FIG. 23B, an indention 145 is then formed in retainingring 142. This indention 145 shortens the retaining ring 142 so thatretaining ends 143 and 144 are drawn closer together. Retaining ring 144exerts a compressive force on cylinder head land 139 and cylinder walland 140 bringing the two lands closer together and applying acompressive force on cylinder chamfered edge 146 which opposes the flatsurface 150 of the cylinder head. The pressure asserted by the retainingring on the chamfered edge 146 forms a seal between the chamfered edge146 and the flat surface of the cylinder head 150. Thus, a metal tometal seal is formed in the combustion chamber between the cylinder edge146 and cylinder head surface 150. The retaining ring and lands form aflange which fits into a female groove 195 formed in engine half 26(FIG. 5).

The combustion pressure between the top of the piston crown and theinterior surface of the combustion chamber formed by the cylinder walland the interior of the cylinder head tends to assert a force on thecavity formed between the chamfered edge 146 and flat surface 150.Pressure within this small area is negligible and not threatening to theintegrity of the seal between the cylinder and cylinder head. Anycombustive force that leaks through the cylinder head seal, if any,exerts a negligible pressure on the gap formed between chamfered edge146 and flat surface 150.

Lubrication System

In the present example of a preferred embodiment of the presentinvention, the engine provides an independent lubrication system foreach engine. Referring to FIG. 4, each engine contains an independentlubrication and cooling system comprising a coolant pump 402, ascavenger pump 404, and a pressure pump 406. When engines areinterconnected, the coolant and lubrication fluids are manifolded 408a,408b and 408c in parallel to each engine so that each engine is suppliedwith lubricating and coolant fluid at the same temperature. Thus, eachengine runs at the same temperature. A plurality of engines connectedtogether to form an extended engine, will have an adequate pumpingsystem because each engine is independently lubricated and cooled. Thereis no need to add additional pumps to an assembly of interconnectedengines other than to manifold the supply to the engines.

The present invention has an advantage over typical engines which supplycoolant and lubricant serially to each engine. Typical engine designsprovide for serial coolant and lubricate distribution. Coolant andlubricate are first run through a first engine before they are runthrough the second, third, fourth, fifth engine, etc. In the presentexample of a preferred embodiment, the coolant and lubricant areprovided in parallel to each engine module so that the coolant andlubricant are supplied to each engine at the same temperature, ratherthan preheating the lubricate and coolant in the first engine beforesending it to the second engine and so on. Thus, the present inventionruns at a lower overall temperature and a more constant temperature. Theinherent adequacy of the independent pumps provided within each engineeliminates the need for an engine manufacturer to install custom pumpingsystems to promulgate various numbers of engines connected together toform a engine.

Valves

As shown in FIG. 24, in the present example of a preferred embodiment ofthe present invention, three intake valves 152 and three exhaust valves153 are provided per cylinder head 154. Spark plugs 156 are shown inFIG. 24 for reference. The use of six valves, combined with the low RPMof the engine enables the cylinder head of the present invention toperform using a very small opening under the valve. As valves are liftedonly a short distance and are elliptical or flattened ports to inducetangential gas flow. Six valves generate a large contact area relativeto the overall valve mass and area. Thus, the design of the valves inthe present invention enable rapid heat transfer from the valve to thehead.

Turning now to FIGS. 25 and 26, the configuration of the ports is suchthat all surfaces may be machined with standard milling bits operatingfrom an axis parallel to the valve axis and parallel to the port axis.The shape and angle of the valve housing is such that small valveopenings enable gas flow which reacts to the valve more as a streamlinerather than as a 90° impediment. The entire head above the combustionchamber is pressurized with cooling oil, thus the valve stem spring andcam follower mechanism are immersed in coolant. The flattened orelliptical ports allow for a short heat path to the coolant, as shown inFIGS. 25 and 26.

FIG. 27 shows the crank case engine half, guide key 82. While an exampleof a preferred embodiment of the present invention has been presented,it is not in tended to limit the spirit or scope of the invention.Variations of the preferred embodiment are possible while remainingwithin the scope of the claimed invention.

What is claimed:
 1. A crankshaft and connecting rod bearing assembly,said assembly comprising:a crankshaft section having a male crank pinsection; a crankshaft section having a female crank pin section whereinthe male crank pin section fits inside of the female crank pin sectionto form a crank pin between the camshaft sections; a first spline formedon the external surface of the male crank pin section; a second splineformed on the interior surface of the female crank pin section whichengages the first spline, whereby the male and female crank pin sectionsare rotationally fixed relative to each other; a bearing having a tab,said tabbed bearing encircling the female crank pin section and adaptedto receive at least one connecting rod end, wherein the at least oneconnecting rod end, tabbed bearing, and male and female crank pinsections rotationally cooperate; and at least one oil aperture in saidtabbed bearing positioned so that the tab of said tabbed bearing keepsthe at least one connecting rod end covering said at least one oilaperture for all rotational positions.
 2. The crankshaft of claim 1wherein the crankshaft diameter is sufficiently large so that itsnatural vibration frequency is greater than the engine power impulsesdelivered to the crankshaft.
 3. The crankshaft and connecting rodbearing assembly of claim 1, wherein a plurality of connecting rod endsare attached to said tabbed bearing and a single crank pin.
 4. Thecrankshaft and connecting rod bearing assembly of claim 3, furthercomprising a plurality of oil apertures, wherein the tab of said tabbedbearing keeps said plurality of oil apertures covered by said pluralityof connecting rod ends.
 5. The crankshaft and connecting rod bearingassembly of claim 1, wherein the at least one oil aperture in saidtabbed bearing allows oil to pass through thereby lubricating saidtabbed bearing, said crank pins and the at least one connecting rod end.6. The crankshaft and connecting rod bearing assembly of claim 1,wherein the at least one connecting rod end requires less than 180degrees of said tabbed bearing surface.